JPS61262222A - Bearing supporter - Google Patents

Abstract

(57) [Summary] This bulletin contains application data before electronic filing, so abstract data is not recorded.

Description

DETAILED DESCRIPTION OF THE INVENTION Field of the Invention This invention relates to a bearing support for a rotary engine;
More specifically, Background of the Invention relates to axially and radially compact radially resilient bearing supports with integral fluid motion capabilities for controlling the vibratory power of gas turbine engines. An engine or prime mover is subjected to vibrational excitation due to the high rotational speed of its rotor and high operating temperature. This has been a particular problem in aircraft gas turbine engines. In such engines, the rotor is often supported from a stationary housing or frame of the engine by means of spring mounts to control rotor vibration. Each spring mount typically includes an anti-friction bearing and a bearing support. The radial spring rate and damping characteristics of the mounting section are important factors in ensuring correct operation of the bearing and shaft system at high speeds of the engine.

In gas turbine engines, the rotor shaft often has several resonant vibration peaks due to rotor speed, rotor profile and some unbalance. As a result, some radial movement of the shaft must be provided in order to obtain acceptable vibration limits for the engine. However, this movement must be controlled and damped to prevent internal damage to the engine due to contact between adjacent stationary and rotating parts. Therefore, the bearing support must allow the shaft to undergo some radial vibrational run-out, but must prevent extreme movement of the shaft. Typically, the preferred spring rate for damping such bearing supports in gas turbine engines is 30,000-150.
The force is about 1,000 pounds.

One conventional type of resilient rotor bearing support consists of an axially extending cantilevered spoke-shaped cylinder or cone to support and cushion the shaft bearing. This support, commonly referred to as a "squirrel cage" bearing support, has been used with good success to support rotor shafts. Additionally, such bearing supports are designed to transfer unwanted resonant frequencies due to vibrations to rotor speeds above and below the normal operating speed of the engine. This minimizes the amount of time the engine is exposed to significant vibrational resonance.

In these types of supports, it is useful to incorporate a damping action into the spring mounting section. Typically, the braking effect is generated by supplying oil to the cavity between the bearing support and the frame of the engine. In some cases, a separate brake assembly with an oil-filled chamber and a brake shim between the bearing support and the frame of the engine are used for this purpose. An example of such a damping assembly is described in U.S. Pat. No. 4,289,360.

Although damped elastomeric bearing support systems have been used with success in controlling the vibrational power of modern gas turbine engines, they are not always acceptable for use in some types of engines. For example, a common squirrel bearing support requires considerable axial space for its long cantilevered spokes. This axial space is not always easily available;
The arrangement of squirrel cage bearing supports can interfere with the engine's preferred gas flow path or add to the length and weight of the engine. Furthermore, while the bearing is held at the ends of the cantilevered axial spokes of the squirrel cage support, misalignment of the bearing support with respect to the engine frame will result in uneven loading of the bearing and The useful life may be shortened.

Finally, it should be noted that cage supports are considerably more expensive due to the numerous precision machining operations required to create the features of the support and its associated frame.

Another type of common bearing support, which is relatively small and low-lying, is a flat single-layer circular spring support with circumferentially alternating inner and outer lobes (spaced pads), usually in the form of a "ring". called "spring". Ring springs are cheaper than squirrel cage bearing supports, but have irregular performance. This is because its spring rate changes with its radial displacement, ie it is a non-linear device. Additionally, there is inherently some sliding contact between the ring lobe and the surrounding engine frame. This sliding contact can cause undesirable fatigue of the lobes, which can change the spring rate of the support. As a result, the resonant frequency of the rotor changes, which tends to lead to vibration problems. Furthermore, the ring lobes apply point loads along the circumference of the bearing, which can induce undesirable stresses and deflections. Finally, although such devices are essentially flat, relatively thin circular strips, they do not provide axial support for the bearings and engine rotor, and therefore cannot handle axial thrust loads. I can't do that.

Therefore, it is radially elastic, can provide bearing support for axial thrust loads, and is low-lying and compact, capable of controlling and damping radial vibrations without sliding near points. There is a demand for bearing supports.

SUMMARY OF THE INVENTION Accordingly, it is an object of the invention to provide a radially resilient bearing support for a rotor shaft subjected to vibrations.

Another object of the invention is to provide a bearing support that is relatively compact in the axial direction.

Another object of the invention is to provide a bearing support with axial stiffness to handle rotor axial thrust loads.

Another object of the invention is to provide a bearing support with increased radial stiffness in a reduced radial dimension.

Another object of the invention is to provide a bearing support that acts to resiliently support the bearing and provides integral internal damping without the need for a separate sliding assembly.

Briefly, the objects set forth above, as well as others that will become apparent from the following description, are achieved by the present invention providing an improved bearing support for a rotor shaft. A support includes an elongated arcuate circumferential spring element connected between the outer and inner shells for mounting the bearing between the engine frame and the rotor shaft. Circumferential springs are radially resilient and axially rigid.

This support is used to control radial vibrations of the engine rotor shaft through resonant frequencies, as well as to control axial movement of the rotor relative to the engine frame due to rotor axial thrust loads. It is particularly effective in preventing

The above and other objects and advantages of the present invention will become apparent from the following detailed description of preferred embodiments of the invention with reference to the drawings.

The same reference symbols are used throughout the drawings to refer to the same parts. The drawings are not necessarily to scale, emphasis instead being placed upon illustrating the concepts of the invention.

DETAILED DESCRIPTION In this description, a case will be described in which the present invention is applied to a gas turbine engine, but FIG. 1 only shows a portion of this gas turbine engine, and the explanation here will be limited to that portion. However, it is clear that the invention may be used in all other types of prime movers and engines requiring radially resilient rotor support means to accommodate and control vibrations.

A portion of a gas turbine engine, generally indicated at 10 in FIG. is installed on. The bearing assembly 16 and the shaft 12 surrounded by it are assembled into an overall structure according to one embodiment of the invention.
It is supported inside the frame 14 by a bearing support indicated at 8.

More specifically, the frame 14 has an inner annular web 20 having an axial bore or passageway 22 that receives the bearing assembly 16 and support 18. The passage 22 is dish-shaped at 24 for seating the bearing support 18. The bearing support 18 is held securely in the countersunk hole 24 by an annular locking nut 26 having external threads. The locking nut 26 is turned into an internally threaded extension 28 at the mouth of the countersunk hole 24.

A bearing assembly 16 is attached to a radial shoulder 3 formed on the shaft 12.
2 and has an inner race 30 suitably attached thereto. An inner race 30 supports an array of ball bearings 34 which are held apart by a cage 36. Bearing 1
6 outer races 38 are suitably attached to the inner periphery of the bearing support 18, as will now be described. Also shown in FIG. 1 is a portion of a typical guide W40 used to direct the engine's main airflow into the flow path of the frame 14.

In one embodiment of the invention, the bearing support 18 is mounted on the frame 1 of the engine.
controlling the radial movement of shaft 12 relative to 4; Referring to both FIGS. 1 and 2, bearing support 18
includes a first cylindrical outer shell 42 ) and a second inner shell 44 radially inwardly spaced therefrom and concentric therewith defining an annular space or gap 46 . One or more arcuate leaf springs 48 are arranged in the annular space 46 between the two shells, and these leaf springs extend in the V-circumferential direction (this is why they are referred to as circumferential springs). call), 1st
and the second shells 42, 44 and are spaced apart from each other in a concentric manner and parallel to each other. In the illustrated embodiment of the support, there are four identical circumferential springs 48 distributed circumferentially around the axial center axis of the bearing support. Each circumferential spring 48 has opposite first and second ends 50, 52, the first end 50 being suitably secured to the first shell 42, such as by being integrally formed therewith. . The opposite second end 52 of each circumferential spring 48 is suitably secured to the second shell 44, such as by being integrally formed therewith.

The first and second ends 50,52 of the circumferential spring are respectively
The fastening to the second shell 42.44 can be done by first and second integral support projections 54.56. The protrusions 54,56 are dimensioned such that the circumferential spring 48 is disposed concentrically with the first and second shells 42, 44 and equidistantly spaced V therefrom. Circumferential spring 48
If it has the shape of a spiral, then during vibrational motion, the first
There is a risk that an undesirable relative rotation due to swirling may occur between the second shells 42 and 44, which may cause slippage on the rolling surface of the bearing 16.

Although not required, it is preferred that the shell 42, 44, circumferential spring 48 and support projection 54, 56 are all integrally formed. In some applications, the ends 50.52 of the circumferential spring 48 can be suitably welded to the shells 42,44.

As shown in FIG. 1, the axial width W of each circumferential spring 48 can be much larger than the radial thickness T, so that it can act as an axial thrust bearing. preferable. In order for a gas turbine engine to have an appropriate support capacity for thrust loads, the width to thickness ratio W/
One example is for T to be at least 4:1, although smaller ratios may be used.

Furthermore, since both the first and second ends 50° 52 of the circumferential spring are fixed to the first and second shells 42, 44, respectively, the axial direction of the circumferential spring 48 is The resistance to deflection and torsion increases, but this
This cannot be achieved if the wire is not fixed or if only one end is fixed.

Thus, when the bearing support 18 is installed in the engine 10 shown in FIG. At the same time, the shaft 12 is moved to a neutral position on the longitudinal center axis of the engine. For example, a typical bearing support 18 can generate a restoring force on the order of 50,000 bonds per inch of displacement. On the other hand, the support is very rigid in the axial direction, so that the axial thrust
The shaft 12 is restrained from significant axial movement when transferred onto the bearing assembly 16, support 18, and frame 14.

More specifically, in order to deal with axial thrust,
FIG. 1 shows one embodiment of a suitably formed bearing support 18. A first shell 42 is tightened by a locking nut 26 into a countersink 24 in the frame. The second shell 44 has a radially inwardly extending flange 58 at one end against which one end of the outer race 38 of the bearing abuts. washer 60
is positioned against the second end of the outer race 38, and a conventional retaining ring 62 is suitably secured in a groove 64 formed in the inner surface of the second end of the second shell 44 to hold the outer race 38 against the outer race. 3B is fixed to the second shell 44.

Thus, when an axial thrust load is transferred to the bearing assembly 16 via the shaft 12, this load is transferred via the outer race 38 to the flange 58 and via the circumferential spring 48 to the first shell. to the body 42 and thereafter to the frame 14.

The axial stiffness provided by the bearing support 18 generates a restraining force to cope with axial thrust loads.

Furthermore, the bearing support 18 serves to create an acceptable relatively large radial stiffness within relatively small axial and radial dimensions. This is due in part to the securing of the first and second ends 50,52. Of course, this also depends on the radial thickness and axial width of the circumferential spring 48, which affect the stiffness due to the bending of the circumferential spring 48. The bending stiffness of a certain member is usually known. However, with this invention, extra radial stiffness is provided.

More specifically, FIG. 3 shows another embodiment of the invention in which three circumferentially spaced circumferential springs 48 all have a minimum radial dimension. In order to keep the bearing support in the radial direction, they are arranged circumferentially aligned with each other at one radius R1, so as to obtain a radially coherent bearing support. Additionally, the resultant radial force F that may be applied to the bearing support 18 during operation is shown and is included to illustrate the increased radial stiffness that can be achieved with the present invention. It is. Force F
induces a bending stress at the center location 66 of the circumferential spring 48a, where this force F is directed. Circumferential spring 48a
has a radial spring resistance mainly due to bending. However,
With two circumferential springs 48b placed generally transverse to force F, bending occurs not only at a location 68 extending generally parallel to force F, but also at a location 68 extending generally parallel to force F. Compression of circumferential spring 48b also occurs at 70.

Vertical compression of circumferential spring 48b at position 70 and lateral bending at position 68 causes circumferential spring 48
Increases radial stiffness when compared to bending only at position 66 of a.

Of course, the bearing support 1 can also be used from embodiments other than that shown in FIG.
An increased radial stiffness of B is obtained.

The foregoing is merely to illustrate the manner in which bearing support 18 operates in one embodiment; such operation occurs in other embodiments of the invention as well.

The circumferential spring 4B, in combination with the fixed supports 54, 56 to the shells 42, 44, also provides considerable torsional stiffness. This is mainly due to the compression and bending of the circumferential springs 48 as described above, and also due to the circumferential springs 48 being concentrically arranged. Spiral springs are undesirable because they have relatively low torsional stiffness and can cause unwanted torsional displacement.

As mentioned at the outset, it is desirable for bearing supports for gas turbine engines to dampen resonant vibrations of the engine. The embodiment shown in FIG. 1 incorporates fluid movement means 72 for this purpose. As shown in FIGS. 1 and 2, oil or other damping fluid is suitably supplied to a conduit 74 within the web 20 of the frame. This conduit communicates with one or more small holes 76 in the first shell 42 to pass damping fluid around the annular space 46 between the shells 42 and 44 and around the circumferential spring 48.

As shown in FIG. 1, a portion of the inner end of lock nut 26 is spaced from the side of circumferential spring 48 so that oil can flow freely around circumferential spring 48 and into the second shell. 44
You can enter the space adjacent to . Each locking nut 2
A suitable seal 78 seats in a groove provided in the inner wall of 6 and the inner surface of bore 22. A seal 78 engages an axial flange 80 . 82 of the second shell 44 to contain damping fluid within the bearing support 18 .

Braking is provided by the surface area of circumferential spring 48 cooperating with the braking fluid. In particular, the deflecting circumferential spring 48
locally pumps the braking fluid and dissipates energy. Note that the circumferential spring 48 itself acts as a braking shim as is known in the art. Therefore, the circumferential spring 48 not only provides the necessary radial resiliency and axial stiffness, but also provides a braking function, thus resulting in a relatively simple structure with several functions.

The size and shape of the circumferential spring 48 may be such that a uniform or non-uniform and/or linear or non-linear spring rate is achieved along the circumference of the support 18. This spring rate can also be commonly used to tune the bearing support 18 to a predetermined natural resonant frequency to avoid resonance during engine operation. The spring rate of the bearing support 18 can be varied, for example, by changing the number of circumferential springs 48, their circumferential position, the thickness or length of these elements, or the material from which the bearing support is made. I can do it.

By using three or more circumferential springs of uniform length and thickness and equally spaced them around the circumference of the bearing support 1B, a more uniform and more linear spring rate is achieved. can get. Conversely, a circumferential spring 48 of varying thickness and length
By using non-uniform spacing, non-uniform and/or non-linear predetermined spring rates can be obtained. For example, non-uniformity may be desirable to accommodate dead weight of shaft 12 in bearing 16. For example, a relatively short and/or thick circumferential spring 48 may be placed at and/or near the top of the support 18 so that the weight of the shaft 12 is initially aligned coaxially with the frame 14. By shifting , the initial spring force in the vertical direction can be appropriately introduced.

The circumferential spring 48 can be arranged in various ways within the bearing support 18. 1 and 2 show the bearing support 1
a support 18 with four circumferential springs 48 arranged primarily in one layer between eight first and second shells 42, 44;
It shows. In other engines or applications, it may be desirable to have a greater number of shorter circumferential springs 48 within this layer.

In certain applications characterized by very soft radial compliance, the desired spring rate can be achieved by a single circumferential spring 48 extending over the entire V of the annular space 46 between the two shells 42, 44. There are things you can get. Additionally, three uniformly spaced circumferential springs 48 aligned circumferentially at one radius may be used as shown in the embodiment of FIG.

It is not necessary that all circumferential springs 48 be arranged in one layer between the first and second shells 42,44 of the bearing support 28. For example, FIG. 4 shows another embodiment of the bearing support 84 of the present invention, which has two different concentric layers in the annular space 46 between the first and second shells 42, 44. A circumferential spring 86 is arranged. This example shows four springs 86, each of which is connected to the annular space 4.
6, so that the circumferential springs 86 overlap each other.

Each circumferential spring 86 has first and second arcuate portions 88 .
．． 90, which are preferably interconnected by an integral U-shaped or stepped portion 92, so that the first and second portions 88.90 have different radii R2, R3. A second portion 90 overlaps and is spaced from the first portion 88 of an adjacent circumferential spring 86;
The U-shape (with the exception of the U-shaped portion 92) remains arcuate, ensuring that the first and second circumferential spring portions 88,90 remain concentric with the shells 42, 44. However, the formation of a swirl in the circumferential spring 86 is thus avoided.

Due to the arrangement of the overlapping circumferential springs 86, the bearing support 84 has a more uniform radial spring along the entire circumference of the support due to the longer length of the circumferential springs 86 undergoing bending. It is possible to have a rate.

In other words, no matter which direction in the radial direction the second shell body 44 is bent, the restoring force of the circumferential spring 86 is uniform. Of course, the circumferential spring 86 is longer, reducing the radial stiffness of the bearing support.

When installed in the engine of FIG. 1, the bearing support 84 provides a stronger braking action than the support 28 because the circumferentially longer circumferential spring 86 has less surface area exposed to the braking fluid. This is because it is even larger.

In any embodiment of the invention, for example, as shown in FIG.
By including a thin metal or foil insert 94 in the space between adjacent circumferential springs 86, further braking action can be achieved. Although such a thin insert 94 does not substantially affect the spring rate of the bearing support 86, it increases the surface area exposed to the damping fluid.

Many other variations in the shape of the circumferential springs of the bearing support can be made to achieve the desired spring and damping characteristics for a particular application. For example, FIG. 5 shows another embodiment of a bearing support 96 of the invention having an additional third annular shell 98 spaced radially inwardly from the second shell 44. ing. A plurality of first circumferential springs 48 similar to that shown in FIG.
44 in the same way. Similarly, a plurality of second circumferential springs 100 are secured to the second and third shells 44, 98 within a second annular space or gap 101 therebetween. In this embodiment, the first and second shells 42, 4
There is a layer of four first circumferential springs 48 disposed in a first annular space between four springs. A second layer of three second circumferential springs 100 is arranged within the second annular space between the second and third shells 44, 98. For this reason, the bearing support 96 has a circumferential spring of 48°100
, which has a more uniform spring rate along the circumference of the support.

Furthermore, in the embodiment shown in FIG.
8 extends in a first, counterclockwise circumferential direction from the first shell 42 to the second shell 44 and a second circumferential spring 100
extends in an opposite clockwise second circumferential direction from the second shell 44 to the third shell. This arrangement allows the circumferential spring 4
8,100 can help eliminate any negative effects caused by radial displacement. For example, if the second shell 44 tends to rotate due to deflection, the third shell 9
8 can be counteracted by the tendency to rotate in the opposite direction.

Furthermore, the first circumferential spring 48 can be designed to be relatively stiff for protection against overloads, whereas the second circumferential spring 1° It can be designed to be rigid against the movements of

Furthermore, the first and second circumferential springs 48.100 are arranged in series, resulting in an embodiment having a relatively softer overall spring rate than by simple arithmetic addition. It will be done.

In the bearing supports described so far, the circumferential springs in each layer of the bearing support are generally similar.

Depending on the application, non-uniform spring elements may be used so that the bearing support has different spring rates along the circumference of the bearing assembly. For example, a given support can be designed to have greater stiffness in vertical deflection than in lateral deflection. That is, by increasing the stiffness in the vertical direction in this manner, it is possible to compensate for the effect of gravity on the shaft 12 when the shaft 12 is at rest, as described above. Such non-uniform spring rates can also be used to detune broadband vibrational resonances of the engine rotor.

This bearing support can also be attached to the engine frame 14 in a variety of different conventional ways than that shown in FIG. 1, and the internal shape of the components of the support can be changed accordingly. I can do it.

The embodiment of the invention shown in FIG. 6 shows that conventional bearings other than ball bearings, ie, roller bearing assembly 102, can be used. In this embodiment, the illustrated bearing support 10
It is also shown that the second shell 104 of 6 can also act as an integral outer race for the roller bearing assembly 102. With this type of support, a braking action may not be necessary, so no oil supply or seal is shown in the drawings. In all other respects, this bearing support can be the same as previously described, except that the roller bearings do not transfer axial loads to the frame.

Bearing supports according to other embodiments of the present invention may also be used in conjunction with new materials being introduced into today's engines. For example, engines that utilize ceramic bearings and/or shafts are most useful when high engine temperatures are expected.

7 and 8 are cross-sectional views of the bearing support 108 mounted inside the ball bearing assembly 110 to control radial vibration of the ceramic shaft 112. In this case, the bearing assembly 110 is suitably mounted directly to the engine frame 14 and secured by a locking nut 26 in the same manner as previously described for the support 18 of FIG. The metal outer shell 114 of the support 10g is attached to the inner race 11 of the bearing assembly 110.
6, and the inner shell 118 of the support 108 is secured to the shaft 112 by a locking nut 120 in the same manner as shown in FIG. A circumferential spring 48 is suitably attached to the shells 114,118. The bearing support 108 is thus
12 and is designed to accommodate the temperature imbalance between shaft 112 and bearing assembly 110.

More specifically, the inner shell 118 is interrupted;
Alternatively, at least one axially extending gap 122 is formed. In the illustrated embodiment, four circumferential springs 48 and four circumferentially spaced gaps 122 are shown. These axial gaps 122 are defined by the shaft 112
) Outer shell 114 or bearing assembly 110 of support 108
The inner shell 118 is allowed to expand or contract 6 with the shaft 112 without constraining any of the shafts 112. Therefore, the difference in thermal expansion and contraction between the metal inner shell 118 and the ceramic shaft 112 in contact therewith can be accommodated, and damage to the shaft 112 can be prevented.

Although the inner shell 118 of the support 108 is split, it does not act as a spring because it is in abutting contact with the shaft 112. Only the circumferential spring 48, which is secured to the inner shell 118 and outer shell 114, renders the support 108 radially resilient. Because an independent circumferential spring 48 is used, even if the inner shell 118 is divided, the bearing support 108
There is no adverse effect on the integrity and performance of the

In normal operation of all support embodiments described so far, there is no noticeable fatigue of the sliding surfaces between the support and the engine frame. It is also not expected that intermittent contact will occur between different parts of the support. Therefore, the expected life of the support is not shortened by such fatigue mechanisms or intermittent contact.

As can be clearly seen from FIGS. 1-8, bearing supports embodying the invention are designed to provide favorable spring rates and braking properties for a variety of different rotor and engine frame geometries. Can be easily changed. Furthermore, since the braking action is integral with the circumferential spring of the support, the adjacent oil-filled cavities or shims sometimes used in squirrel cage bearing supports can be omitted. This results in considerable radial space savings and further increases the braking capacity.

More importantly, the long axial cantilever spokes found in conventional squirrel cage bearing supports are eliminated. These spokes take up a considerable amount of space in the axial direction, making the engine longer. For this reason, the weight and cost of the engine <t@7J[l.

The coherent bearing support described here saves engine weight, material and machining that would otherwise be required. The bearing support and the simple frame shape described above result in considerable cost savings compared to conventional square supports.The bearing support of the present invention is easy to assemble; The support part and braking structure of the adjacent engine can be easily removed.

The bearing support body embodying this invention is different from the conventional one-length if
It generates a more uniform restoring force than the conventional one. - The long spring applies a radial point load on the bearing race at each inner lobe, assuming the outer race is installed and zero. In contrast, the bearing support of the present invention generally has no variation in circumferential (or radial) loads. It is most valuable to use thin bearing races, such as roller bearings with roller bearings (the opposite of roller bearings 102).In such bearings, uneven radial loads will cause the outer race to flex and cause the rollers to flex. This creates uneven loading between the races, which is a contributing factor to bearing failure.

Although the invention has been described with particular reference to preferred embodiments, those skilled in the art will recognize that various modifications can be made within the scope of the invention as defined by the claims. For example, the bearing support can be adapted to a variety of different types of bearing assemblies for use in various types of prime movers, engines, and machines. Additionally, the means for supplying damping fluid to the support can also vary depending on the ease of access to the fluid source.

[Brief explanation of the drawing]

FIG. 1 is an axial sectional view of a portion of a gas turbine engine showing a bearing support according to an embodiment of the present invention, and FIG. 2 is an enlarged sectional view taken along section line 2-2 in FIG. The bearing support of the engine is shown in more detail. FIG. 3 is a cross-sectional view of an alternative embodiment of the bearing support for use in the engine of FIG. 1, using three circumferentially aligned springs. FIG. 4 is a cross-sectional view of a further embodiment of the bearing support, which uses overlapping circumferential springs with foil inserts to enhance the braking action. FIG. 5 is a cross-sectional view of a further embodiment of a bearing support for use in the engine of FIG. 1, having three concentric shells with two sets of circumferential springs formed therebetween. It is located within the inner and outer annular gaps. 6th
The figure is a partial cross-sectional view of an alternative embodiment of a bearing support for use in the engine of FIG. 1, showing the roller bearing assembly and its radially outer sawn-mounted bearing support. Figure 7 is a cross-sectional view similar to Figure 2 of a further embodiment of the bearing support, the support being angled radially inwardly than the bearing assembly and supporting a ceramic shaft; . FIG. 8 is a cross-sectional view taken along section line 8--8 in FIG. 7. Explanation of main symbols 42.44: Shell body 48 Two circumferential springs 50.52: End

Claims (1)

[Scope of Claims] 1) an annular first shell; a second annular shell coaxial with and separated from the first shell, and creating an annular space therebetween; at least one elongated arcuate circumferential spring disposed within the annular space and substantially concentric with the first and second shells, the first and second shells on either side of the circumferential spring; A bearing support having second ends secured to said first and second shells, respectively. 2) In the bearing support described in claim 1),
A bearing support having three circumferentially spaced circumferential springs. 3) In the bearing support described in claim 2),
A bearing support in which the three circumferential springs are circumferentially aligned at one radius. 4) In the bearing support described in claim 2),
A bearing support having bearing means disposed radially inward from the second shell, the first shell being able to be attached to a frame of an engine. 5) In the bearing support described in claim 2),
A bearing support having bearing means disposed radially outside the first shell and radially inside the engine frame, the second shell being able to be attached to the engine shaft. 6) In the bearing support described in claim 2),
A bearing support wherein the circumferential spring is sized and shaped to produce a substantially uniform radial restoring force along the circumference of the support. 7) In the bearing support described in claim 2),
A bearing support, wherein the size and shape of the circumferential spring is such that it produces a radial restoring force that is not uniform along the circumference of the support. 8) In the bearing support described in claim 2),
A bearing having a gap passing through the second shell in the axial direction and capable of coping with differential thermal expansion and contraction between the annular member and the annular member disposed in butt contact with the second shell. support. 9) In the bearing support described in claim 8),
A bearing support in which the annular member comprises a ceramic shaft disposed radially inside and coaxially with the second shell. 10) In the bearing support according to claim 2), each circumferential spring has a radial thickness and an axial width, the width being greater than the thickness and an axial width. A bearing support that can function as a thrust bearing. 11) In the bearing support according to claim 10), the width to thickness ratio of the circumferential spring is at least 4:
1. 12) In the bearing support according to claim 2), each circumferential spring is composed of first and second parts interconnected by a U-shaped part; A bearing support in which the second circumferential spring portions have different radii and the second portion of the circumferential spring overlaps the first portion of an adjacent circumferential spring. 13) In the bearing support according to claim 12), braking means includes means for acting to pass oil into the annular space between the overlapping first and second portions of the circumferential spring. and for this purpose the circumferential spring has a surface area effective to generate a radial spring force on the bearing support and to cooperate with the oil to damp vibrations. support. 14) In the bearing support according to claim 13), a foil insert is arranged in the annular space between the circumferential springs and has an additional surface area for increasing damping. A bearing support having: 15) In the bearing support according to claim 1), a plurality of the first circumferential springs are connected to the first and second circumferential springs.
a third annular shell is disposed between said second shells and is disposed radially spaced from said second shell and coaxial therewith defining a second annular space therebetween; A bearing support in which a second circumferential spring is secured to said second and third annular shells. 16) A bearing support according to claim 15) having four first circumferential springs and three second circumferential springs. 17) The bearing support according to claim 15), wherein the first circumferential spring extends in a first circumferential direction from the first shell to the second shell; A bearing support in which the second circumferential spring extends in an opposite second circumferential direction from the second shell to the third shell. 18) In a bearing support for a gas turbine engine, a first annular shell and a second annular shell coaxial with but spaced apart from the first annular shell creating an annular space therebetween. a shell, and at least three shells disposed within the annular space, circumferentially spaced apart and substantially concentric with the first and second shells.
elongated arcuate circumferential springs, each circumferential spring having opposing first and second ends secured to the first and second shells, respectively; The directional spring has a radial thickness and an axial width, the width being greater than the thickness to enable it to act as an axial thrust bearing; damping means including means for passing oil through the annular space of the bearing so that the circumferential spring generates a radial spring force on the bearing support and cooperates with the oil to damp vibrations. A bearing support having an effective surface area to 19) A bearing support according to claim 18), wherein each circumferential spring has first and second portions interconnected by a U-shaped portion; A bearing support wherein portions of the circumferential spring have different radii such that a second portion of the circumferential spring overlaps a first portion of an adjacent circumferential spring.